The present invention relates to balance mechanisms for rotating machinery, particularly balance shafts for multi-cylinder internal combustion engines which exhibit vertical shaking forces.
Balance shafts are commonly used to reduce or cancel shaking forces and/or vibrations which result from residual imbalances inherent in the design architecture of machinery with rotating and/or reciprocating parts, or mechanisms, such as motors. These balance shafts are sometimes called xe2x80x9ccounterbalancexe2x80x9d shafts.
Balance shafts are particularly valuable when operator or passenger comfort and freedom from noise and vibration related fatigue or distraction are desired, as in the case of motor vehicles such as automobiles, motorcycles, and the like. It is also advantageous to minimize vibration from the standpoint of equipment reliability. Where vibrations are reduced, the size, mass, and/or complexity of the mounting structures can often be reliably reduced, thus potentially reducing costs.
With multi-cylinder motor vehicle engines, the inline four-cylinder engine configuration is favored in much of automotive and industrial use today due to its inherent packaging space, manufacturing cost, and fuel consumption efficiencies. These engines benefit from Lanchester-type balance shafts, which can cancel nearly all of the inherent twice-per-revolution shaking forces produced by this otherwise mass-balanced architecture.
Lanchester-type balance shafts for these inline four-cylinder engines are paired to rotate in opposite directions at twice engine speed. The two balance shafts are timed to cancel each other""s lateral forces while opposing the vertical xe2x80x9csecondary shaking forcesxe2x80x9d that result from connecting rod tilt, causing piston motion to depart from sinusoidal or xe2x80x9csimple harmonicxe2x80x9d motion in the midstroke region centered about 90 degrees before and after xe2x80x9ctop dead center.xe2x80x9d Each shaft produces a single, or xe2x80x9cstaticxe2x80x9d rotating unbalanced or centrifugal force, which taken together with its mating shaft""s rotating unbalanced force, produces a resultant vertical shaking force which most effectively is located centrally among the bank of cylinders so as to be coincident with the engine""s resultant shaking force. Static unbalance-type shafts of this general type are shown, for example, in U.S. Pat. No. 5,857,388.
Helical gears are often employed as the means of maintaining orientation xe2x80x9ctimingxe2x80x9d between Lanchester-type balance shafts because of their potential to represent the best value choice for engineering priorities such as durability, wear resistance, noise emissions, cost, packageability, mass, power consumption and the like. This potential, especially in the often critical case of noise emissions, is highly dependent upon actualization of the gearset""s theoretically possible total contact ratio (TCR). This, in turn, is much more highly dependent upon freedom from axial misalignment of gear tooth geometries, i.e., gear tilting or twisting with respect to its mate, throughout the operating speed range, than are alternative timing drive means such as chains or toothed belts.
Additionally, it is often advantageous to use a driven balance shaft to drive another component, such as an oil pump, as taught in U.S. Pat. No. 5,918,573 entitled xe2x80x9cEnergy Efficient Fluid Pump,xe2x80x9d to provide a synergistic noise control benefit. In this case, also, the stability of the driving means"" axis alignment throughout all operating speeds can play a critical role with regards to many of the same engineering priorities mentioned earlier. Given the elastic compliance inherent to structural materials, operating shape changes will accompany the load changes that occur as a balance shaft is operated at various speeds. Adding material to increase the section modulus of structural members, such as connector portions of balance shafts, carries a mass (and often cost) penalty, and can carry a fuel consumption penalty and/or packaging space penalty as well, and yet can still not succeed in providing for engineering target value axis alignment control.
Accordingly, there exists an advantage in the ability to achieve targeted axis alignment control for balance shaft drive means such as gears and extensions throughout the operating speed range without incurring additional mass, cost, fuel consumption or packaging space penalty.
With these motivations in mind, an example of a typical balance shaft for an inline four-cylinder engine is shown in a purely static condition in FIG. 3. This prior art balance shaft 10 includes counterweights 12, 14 located adjacent to either side of (or xe2x80x9cstraddlingxe2x80x9d) the principal balance shaft journal 16 so as to apply their composite centrifugal loading to the journal bearing 16, thus transmitting the centrifugal loads to the engine""s structure. This arrangement, with the diameter of the bearing journal being smaller than the effective diameter of the counterweight (as defined by the locus of the counterweight""s largest effective radius as the shaft rotates), represents the combination of power-consuming bearing friction and space-consuming unbalance mass best able to maximize fuel consumption, mass, and packaging space efficiencies in most cases. The bending stiffness of the ideally sized journal is typically lower, however, than that of a larger diameter, suboptimal (in terms of friction and heat generation) journal configurations. Journal bending stiffness typically plays a significant role in operating shape stability versus operating speed, yet can be utilized to advantage without penalty using the inventive strategy disclosed herein.
Referring to FIG. 4, which illustrates the balance shaft 10 with high speed operating shape that is greatly exaggerated for clarity, the counterweights 12, 14 of this prior art configuration are xe2x80x9cbalanced,xe2x80x9d in terms of their bending moments, the product of their respective centrifugal force magnitudes and locations, about the midpoint of the length of the principal journal 16 such that their composite resultant force is located at the midpoint 18 of the journal 16 when the shaft 10 is spun about its axis of rotation CL, 30. By locating the composite resultant force at the journal midpoint 18 in this fashion, the balance shaft""s output loads are typically located at the midpoint of the engine""s cylinder array, which is typically the midpoint of its central bulkhead""s axial length. This is accomplished by distributing the counterweight masses such that their xe2x80x9cmomentsxe2x80x9d of unbalance sum to zero at the midpoint 18 of the journal 16, which is the targeted Effective Plane of Static Unbalance (xe2x80x9cEPSUBxe2x80x9d) location. This is discussed in more detail in U.S. Pat. No. 6,237,442, entitled xe2x80x9cHigh Value Static Unbalance-Type Balance Shafts.xe2x80x9d
In other words, the product of unbalance magnitude and its effective distance from the EPSUB is the same for each counterweight 12, 14, with the xe2x80x9ceffective distancexe2x80x9d being measured from the center of gravity (CG) of each unbalance mass, respectively. These opposingly xe2x80x9cbalancedxe2x80x9d bending moments act to bend the shaft structure principally in the region of the principal journal 16, tending to deflect both ends of the shaft 10 away from its axis of rotation CL, 30 toward the CGs of the unbalance masses at elevated rotational speeds. The end result of this bending acts to result in a tilt, or xe2x80x9cwobblexe2x80x9d (lateral runout) of the drive gears 20, 22 whose quiet operation depends heavily on axis alignment with respect to each other, i.e., the avoidance of such tilt. This unwanted gear tilt is generally represented by reference number 38.
The mechanism which results in this tilt is an axial shift in the effective plane of support to the principal journal 16 provided by the sleeve bearing""s oil film, as generally indicated by reference number 24. This axial shift is a result of any tilting, or misalignment (e.g., due to high speed bending), of the principal journal 16 with respect to its bearing sleeve. Since the balance shaft""s other xe2x80x9coutriggerxe2x80x9d bearing 26 resists motion away from the axis of rotation CL, 30, most of this high speed bending about the principal journal 16 is manifested in radial deflection of the shaft""s overhung end 28, as generally indicated by reference number 36.
This deflection of the overhung end 28 tends to cause xe2x80x9cedge loadingxe2x80x9d of the bearing sleeve due to the tilting of the principal journal 16, which thereby shifts the oil film""s effective xe2x80x9ccenter of pressurexe2x80x9d or OFCOP 24 towards the overhung end 28 where the oil film is hydrodynamically driven or xe2x80x9cwedged,xe2x80x9d into the smallest gap, or xe2x80x9cminimum film thickness.xe2x80x9d With this oil film xe2x80x9ccenter of pressurexe2x80x9d 24 acting as the effective fulcrum where loads concentrate, its relocation towards the overhung end 28 results in the xe2x80x9ccradledxe2x80x9d (between bearings) counterweight 12 having a greater unbalance moment about this fulcrum than the overhung counterweight 14, causing residual loading to be borne by the outrigger bearing 26. This increased moment, resisted by the residual loading on the outrigger journal 26, acts to bend the shaft or connector portion 32, which extends between the bearings 16, 26. The shaft 10 bends during operation in somewhat catenary-appearing fashion, thus tilting the axis of an originally true-running gear 20 mounted adjacent to the outrigger bearing 26 as the shaft""s static centerline 34 deforms from its unloaded straightness, as evidenced by the angle represented by reference number 36.
It is therefore desirable to provide a static unbalance-type balance shaft for a four-cylinder engine that provides operating shape stability versus operating speed for associated drive gears and/or shaft driven extensions without the addition of unnecessary mass or inertia. It is also desirable to provide a static unbalance-type balance shaft that avoids the requirement for indirect, or inefficient, structural load paths to the engine""s cylinder block.
It is therefore an object of the present invention to provide static unbalance-type balance shafts for multi-cylinder internal combustion engines which maintain axis alignment of drive gears or drive extensions at all operating speeds such that the drive gears can be designed to operate with maximum performance and cost efficiency.
It is a further object of the present invention to provide static unbalance-type balance shafts with increased packageability, mass, inertia, and cost benefits.
It is another object of the present invention to enable static unbalance-type balance shafts with two efficiently sized journals to yield equivalent or superior gear axis alignment stability at elevated rotational speeds as compared to balance shafts having a larger number of journals.
It is yet a further object of the present invention to provide static unbalance-type balance shafts that equalize the unbalance moments about the desired Effective Plane of Static Unbalance with minimal overhung or cantilevered counterweight unbalance magnitude and thus minimal journal bending moment.
It is still another object of the present invention to provide a static unbalance-type balance shaft with improved operating shape stability versus rotational speed, for alignment of drive gears and/or extension shafts during rotation, without the need for additional mass to increase structural stiffness.
In accordance with the above and other objects of the present invention, a static unbalance-type balance shaft is provided. The balance shaft is intended to help cancel any unbalance force of an engine, particularly when driven and timed in mirror-fashion with a mating shaft, e.g. Lanchester. The balance shaft has an axis of rotation and includes a first bearing journal located adjacent a first end of the shaft. The balance shaft has a second bearing journal located adjacent a second end of the shaft. The first bearing journal has a first counterweight positioned adjacent one side thereof and a second counterweight positioned adjacent the other side thereof. The first counterweight and the second counterweight are offset from the balance shaft axis of rotation in the same direction. The balance shaft has a third counterweight located adjacent the second bearing journal. The third counterweight is smaller in unbalance magnitude than either the first or second counterweights.
Alternative but functionally equivalent structures include the substitution or supplementation of outrigger counterweight unbalance through the inclusion of deliberately unbalanced gears, the deliberate use of asymmetrical material located in the connector (or between counterweights) portion of the shaft, and/or the extension of the true running end of the inventive three-counterweight shaft for such considerable distance, such as to or through an adjacent engine bulkhead region wherein a third bearing journal may be needed to withstand radial loads from, e.g., drive chain means. In this latter case, the original xe2x80x9coutriggerxe2x80x9d journal (now the center of the three journals) would typically be retained to control gearset center distance, even if it were not required from the standpoint of gear tilt control.
These and other features and advantages of the present invention will become apparent from the following description of the invention when viewed in accordance with the accompanying drawings and appended claims.